The present invention relates to hydraulic lash adjusters for internal combustion engines.
Automobile engines use only a small fraction of their rated power during most of the running time. It is known that increased fuel economy can be achieved by reducing the air pumping losses to the engine cylinder during steady state running, if in particular, some of the engine cylinders are deactivated while the other cylinders are kept active.
There are several ways to achieve this cylinder deactivation. One way is a collapsible hydraulic lash adjuster, whereby engine valves are selectively deactivated. A typical hydraulic lash adjuster is a very simple device, consisting basically of a hydraulic cylinder and piston assembly, mounted either in series or in parallel with the valve train. The working chamber of this lash adjuster is connected to the engine lube oil circuit via a one-way check valve. During the time while the engine valve stays open, the valve closing forces are supported exclusively by the column of lube oil trapped in the chamber. Because of the increased pressure level, some of the initial lube oil charge leaks out, shortening the valve train length and insuring proper seating of the valve. Once the valve is seated and the valve closing force is supported by the valve seat, the pressure in the chamber drops. The gap created by the leakage is then quickly refilled via the one-way (no return) valve from the lube oil circuit. By elimination of the gap there is no significant acoustic noise generated and any seat wear is compensated.
During the engine valve active cycle (valve open), collapsing of the lash adjuster piston assembly is prevented by a lateral latching pin, locked in a corresponding bore of the outer sleeve. During the de-activation cycle, lube oil from a secondary circuit pushes the latching pin out of engagement (against a reset spring) and the lash adjuster carrier, from that point on, will not be able to support the valve train forces and the valve will remain closed (and by that de-activated). The motion of the valve train generated by the cam is instead absorbed by the spring(s) mounted below the lash adjuster carrier.
The disadvantages of this design are, first, difficulties associated with the latching pin to find its target bore during the very short time available for re-activation (especially critical at higher speed) and, secondly, the high bending (shearing) forces the pin and its retaining bore are exposed to.
According to the present invention, the hydraulic lash adjuster is modified so that, upon receipt of a valve deactivation signal, the lash adjuster stop limit more reliably and consistently changes from a hard stop to a soft stop. As a result, the excess force stored in the valve closure spring, displaces the lash adjuster through the soft stop such that the tappet pivot point on the lash adjuster is also displaced to a position where the overhead cam acts with reduced force on the roller finger. Thus, the valve does not open during any portion of the cam shaft rotation. Upon denergization of the lash adjuster, the pivot point for the finger arm returns to the normal position, the lash adjuster encounters a hard stop, and the cam can overcome the valve closure spring to open the valve according to the cam timing.
In essence, a generally conventional lash adjuster is modified by incorporating a coaxially oriented hydraulic control piston assembly within the guide body. The control piston normally fixes latch means, such a plurality of hard spheres, in multiple detents loaded in compression with the other components, to provide a rigid stop, but when the control piston is hydraulically pressurized, the detents are overcome and the piston assembly provides a resilient or soft stop that accommodates extended displacement (retraction) of the lash adjuster within the guide. The hydraulic actuation is preferably implemented with a three-way solenoid valve or the like, for controlling high-pressure oil to a gallery and associated inlet ports for the control piston assembly. In the typical implementation of the invention, the piston need have only two operational positions-denergized to establish the detent or hard stop condition, or fully energized to establish the valve deactivation position.
With all of preferably four detents in quadrant symmetry and associated components in compression, side loading is avoided. Moreover, with the present invention, backlash is also avoided.
More particularly, during high power operation (engine valve active) a substantially cylindrical lash adjusting tappet insert is supported by a ring of balls located in one or more cross holes in the lower portion of the tappet body, engaging with a corresponding annular groove in the guide body bore. The hydraulic control piston is located on the centerline of the tappet body and, energized by its own return spring, keeps the balls spread apart so long as there is no pressurized oil present in the control gallery or chamber. All components supporting the valve actuation reaction forces are loaded in compression in a similar way to a ball bearing, which is very advantageous as far as wear and life expectancy are concerned.
Once the pressurized lube oil is switched on, hydraulic force will overpower the control piston return spring force and move the control piston in the downward direction, allowing the balls to slide down the ramp of the annular groove and by that move towards the center and release the tappet. In this position, the only force trying to push the tappet up is the force of the tappet return spring (deactivation spring) located in the lower portion of the tappet, which is much smaller than the force necessary for valve actuation and by that preventing opening of the associated engine valve.
In order to reduce the contact stress (Hertzian stress) at the most critical point, the upper portion of this hydraulic control piston is preferably shaped somewhat like a compound pyramid, defining four symmetric pairs of upper and lower ramps. Upon activation of the control piston, the balls move from support at the lower ramps to support at the upper ramps. At the same loads the contact stress between a ball and a flat is much smaller than the contact stress between a ball and a cylinder. Also the included angle of both ramps (lower and upper) can be designed in such a way as to minimize resulting reaction force at the ball/ramp interface. In a similar way the locking surfaces (lower ramp) of the control piston can have a small included (self-locking) angle to eliminate backlash during the valve active (balls engaged) period.